It is known that radial bearings of a rotating shaft require an accurate co-centric alignment. However, the use of a precision bearing component cannot always ensure such a requirement in some applications. It is particularly true in a situation in which a rotating shaft is rotatably supported by three or more radial bearings. The alignment errors inevitably cause small radial displacement of the rotating shaft and, as a result, friction forces between the rotational components and stationary components are produced. Consequently, premature bearing failure may occur.
In consideration of dynamic requirements, the use of precision bearing components together with an effective system of a bearing lubrication is required for a rotating shaft driven at relatively high rotational speeds, such as a turbine shaft. When operating at a high rotational speed, some rotating shafts emit audible whine or noise, primarily as a result of high frequency radial shaft excursion caused by the imbalance of rotating components. In many rotary journal applications, it is not possible to alter the bearing span, or the geometry of the journal or the characteristics of the bearing to provide a resonant-free design in the bearing and its housing which is complimentary to the desired or necessary speed range of the journal. Accordingly, it is necessary to provide a bearing housing assembly which will eliminate the rotating shaft excursion and vibration to prevent premature bearing failure, expensive maintenance and repair.
Therefore, a soft bearing structure is desirable for both static and dynamic applications to accommodate lateral displacement of a bearing and damping lateral vibration of the bearing with respect to a static support structure.
Efforts have been developed in the industry to provide a soft bearing support structure. For example, U.S. Pat. No. 4,457,667 which issued to Seibert et al. on Jul. 3, 1984 is entitled "VISCOUS DAMPER WITH ROTOR CENTERING MEANS". Seibert discloses, in this patent, a method to utilize a centering spring for a viscous damped bearing of a aircraft gas turbine engine so that the bearing rotor of the shaft does not bottom against the bottom wall of the reservoir of the viscous damper during aircraft manoeuvres. The spring rate of the centering spring is optimized so as to minimize relative motion of rotor to case while maintaining the required amount of motion within the viscous damper. The centering spring comprises a group of spring rods which are assembled individually to support the viscous damper. Therefore, time and labour are extensively required to assemble this structure. Another problem is that the bearing is actually supported by both the viscous damper and the centering spring and, therefore, the flexibility or softness of the bearing is a result of the combination of the two support structures. It is difficult to precisely adjust the flexibility or softness of the bearing to meet a predetermined requirement.
As another example, in U.S. Pat. No. 3,979,155 which issued to Sood et al. on Sept. 7, 1976 and is entitled "FLEXIBLE DAMPERED BEARING SUPPORT", Sood discloses a flexible damped bearing support which comprises a moveable member for supporting the bearing within a stationary frame, a flexible spring arrangement for suspending the moveable member from the frame and centering the bearing with respect to the shaft, a uniform fluid film squeezed into the post between the frame and the moveable housing for dampering the bearing response and at least one helper spring positioned below the center line of the shaft being arranged to act upon the moveable member for supporting the dead-weight of the rotor structure. Such a construction works in a principle similar to that taught by Seibert in U.S. Pat. No. 4,457,667, except for the additional helper spring for supporting the dead-weight of the rotor structure. The bearing support disclosed by Sood also has as a problem that it is difficult to precisely calculate and adjust the flexibility or softness of the bearing support. A group of spring beams is provided rather than individual spring beams is formed integrally with the moveable member which is a part of the damper. Therefore, the moveable member has to be replaced if the spring beams are to be changed for a purpose of adjustment of the flexibility. The spring beams extend axially and are supported to the frame spaced apart from the moveable member axially rather than radially, which requires more axial space for a desired flexibility.
A further example is shown in U.S. Pat. No. 5,161,940 which issued to Newland on Nov. 10, 1992 and is entitled "ANNULAR SUPPORT". In this patent, Newland discloses an annular support which includes first and second axially spaced axial support rings connected by a plurality of elongated transfer members spaced equally about the circumference of the rings. The transfer members are configured so as to have first and second cross-sectional bending movement inertia, with the second moment of inertia being at least one order of magnitude greater than the first moment of inertia. The transfer members are oriented with the axis defining the first moment of inertia of each transferred member lying tangent to the circumference of the annular support rings. Thus the transfer beams are relatively easily flexed radially as one ring or the other moves radially, but are relatively inflexible for attempted lateral or circumferential displacement in the local circumferential plane. Due to the preferential orientation of the moment of inertia in the transfer members, the annular support disclosed in this patent prevents most relevant movement between the first and second rings while permitting uniformed differential radial expansion typically caused by differential thermal conditions. Therefore, it is apparent that the annular support is not applicable to accommodate lateral displacement and damp vibrations.
There exists a need for a soft bearing support structure to accommodate lateral displacement and to dampen lateral vibration, overcoming the shortcoming of the prior art.